Constant velocity joints connect shafts such that the speeds of the shafts connected by the joint are absolutely equal at every instant throughout each revolution. This distinguishes constant velocity joints from simple universal joints. Specifically, if one of the shafts connected by a universal joint is revolving at an absolutely constant speed, then the other shaft will revolve at a speed that is, during two parts of each revolution, slightly greater and, during the other two parts of the revolution, slightly less than the constant speed of the first shaft. The magnitude of this fluctuation in speed increases as the angle between the axes of the two shafts increases. This disadvantage becomes of practical importance in applications requiring constant velocity such as front wheel driven vehicles and in the drives to independently sprung wheels where the angles between the shafts may be as large as 40.degree..
It is known that the speed variation problem can be solved by using two universal joints in series. If the joints are properly arranged, the irregularity introduced by one joint will be cancelled out by the equal and opposite irregularity introduced by the second joint. Constant velocity joints include such double universal joints as well as any joint in which the speeds of the shafts connected by the joint are absolutely equal at every instant throughout each revolution. Typically a constant velocity joint includes a shaft with a universal-type coupling at each end. This arrangement is sometimes referred to as a constant velocity shaft.
In a front wheel drive vehicle, constant velocity drive shafts are always used in pairs. One shaft is located on the left (driver) side of the vehicle and the other is placed on the right (passenger) side. Each shaft has an inboard or plunge coupling that connects the constant velocity shaft to the engine/transaxle and an outboard or fixed coupling that connects the shaft to a left or right wheel. The inboard and outboard couplings and shaft together comprise a constant velocity joint or drive shaft which couples the engine/transaxle shaft to the wheel shaft. In operation, the outboard coupling turns with the wheel around a "fixed" center, while the inboard coupling "telescopes" or plunges and turns at an angle sufficient to allow required movement of the car's suspension system.
Each coupling must be capable of pivoting at least about two transverse axes to the extent required by the specific application. For example, a compact constant velocity joint that provides power to the wheels typically must operate at angles of 40.degree. or more to meet the car's requirements for steering and suspension movements. Thus, each end of the joint must be able to move at least 20.degree..
Various constant velocity joints have been developed for use in motor vehicles. These include the Tracta joint manufactured in England by Bendix Limited, the so called Weiss joint manufactured in America by Bendix Products Corporation and a joint developed by Birfield Transmissions Limited. Today, there are two basic outboard joint designs and three basic inboard joint designs commonly in use.
The two basic outboard front wheel drive couplings are the Rzeppa and the fixed tripod design. The Rzeppa design includes a cage, inner and outer races and a matched set of six balls guided by the cage. The fixed tripod design includes a three legged cross or trunnion fixed in a housing, a shaft end having a tulip shape, and tracks of circular cross-section to match the rollers.
The three basic types of inboard front wheel drive couplings are the cross groove design, the double offset design and the tripod-plunge design. The cross groove design includes a cage, angled inner and outer races, and a matched set of six balls, guided by the cage for movement in the races. The double offset design is similar to the Rzeppa design and includes a cage, inner and outer races having grooves formed therein, and six balls guided by the cage. The tripod plunge design includes a three legged cross or trunnion and a bearing assembly fixed in place on a splined shaft. The assembly slides in a grooved tulip shaped housing.
One of the basic requirements of the inboard plunging joint or coupling is that it must be able to transmit torque into the wheel axle. The previously mentioned inboard plunging couplings have performed satisfactorily in small cars with relatively low torque engines. However, such couplings have not performed well when applied to larger cars with higher torque engines. Accordingly, there have been attempts to increase the torque carrying capacity of known inboard plunging joints.
One inboard plunging Joint: designed by General Motors to minimize ride disturbance induced by high angulation under high torque, known as "shudder", is shown in FIGS. 1 and 1A. This joint is called the S-plan joint and is said to provide shudderless operation.
As shown in FIG. 1, the S-plan joint is a modified version of the tripod plunge design inboard joint. The S-plan joint typically includes a drive canister or housing 10 having axial grooves formed therein, a trunnion 30' having a splined shaft receiving opening and three legs, a bearing assembly 60 supporting each leg in an axial groove and a flexible boot assembly including a boot 40, sealing ring 41 and clamp 42 for sealing the interior of the joint. Snap rings 6 are provided to retain an engine/transaxle shaft 1 in the splined opening of the trunnion 30'. The principal difference between the S-plan joint and a conventional tripod plunge design PV joint is that the bearing assemblies 60 of the S-plan joint are square so that the torque transmitting surface area is increased significantly. The increased torque carrying capacity of this joint eliminates angulation under high torque (shudder).
The principal disadvantage of the S-plan joint is that the square bearing assemblies 60 responsible for the improved torque capacity results are very intricate and expensive. As best shown in FIG. 1A, each square bearing assembly 60 includes an outer housing 62, outer races 61 and inner races 64 and a series of tiny needle bearings 63 between the outer race 61 and inner race 64. This complex multi-part structure is quite expensive both in terms of cost of the parts and assembly time. This expense is significant since each vehicle requires six such bearing assemblies.
Thus, there is a need for an inexpensive, easily assembled inboard plunging coupling capable of transmitting high torque.
The present invention also relates to the use of bearing sleeves instead of rolling element bearings.
This application relates, in part, to the use of sleeve bearings which can be used instead of expensive ball bearings. The principal limitation in a sleeve bearing's performance is the so-called PV limit. For instance, high edge loading causes a sleeve bearing to reach its PV limit. PV is the product of load or pressure (P) and sliding velocity (V). A sleeve bearing subjected to increasing PV loading will eventually reach a point of failure known as the PV limit. The failure point is usually manifested by an abrupt increase in the wear rate of the bearing material.
As long as the mechanical strength of the bearing material is not exceeded, the temperature of the bearing surface is generally the most important factor in determining PV limit. Therefore, anything that affects surface temperature--coefficient of friction, thermal conductivity, lubrication, ambient temperature, running clearance, hardness and surface finish of mating materials--will also affect the PV limit of the bearing.
Thus, the first step in selecting and evaluating a sleeve bearing is determining the PV limit required by the intended application. It is usually prudent to allow a generous safety margin in determining PV limits, because real operating conditions often are more rigorous than experimental conditions.
Determining the PV requirements of any application is a three step process. First, the static loading per unit area (P) that the bearing must withstand in operation must be determined. For journal bearing configurations, the calculation is as follows: EQU P=W/(d.times.b)
P=pressure, psi (kg/cm.sup.2) PA1 W=static load, lb (kg) PA1 d=bearing surface ID, in. (cm) PA1 b=bearing length, in. (cm) PA1 V=surface velocity, in/rain (cm/min) PA1 N=speed of rotation, rpm of cycles/min PA1 d=bearing surface ID, in. (cm)
Pressure (P) should not exceed certain maximum values at room temperature. These can be derived from a table of allowable static bearing pressure for most known materials. Next, the velocity (V) of the bearing relative to the mating surface must be calculated. For a journal bearing experiencing continuous rotation, as opposed to oscillatory motion, velocity is calculated as follows: EQU V=(dN)
where:
Finally, calculate PV as follows: EQU PV(psi-ft/min)=P(psi).times.V(in/min)12
or, in metric units: EQU PV(kg/cm.sup.2 -m/sec)=P(kg/cm.sup.2).times.V(cm/min)/6000
The PV limits of unlubricated bearing materials are generally available from the manufacturer of the material or from technical literature. Since PV limits for any material vary with different combinations of pressure and velocity as well as with other test conditions, it is prudent to consult the manufacturer for detailed information.
One material which is particularly well suited to bearing applications is the polyimide thermoset material sold by Dupont under the trademark VESPEL.TM.. Properly lubricated VESPEL.TM. parts can withstand approximately 1 million psi-ft/min.